Recirculating type once-through steam generator



Se t. 16, 1969 w. H. CLAYTON, JR

RECIRCULATING TYPE ONCE-THROUGH STEAM GENERATOR 3 Sheets-Sheet. 1

Filed Dec. 27, 1967 k M Ne s WW INVENTOR. WILL/AM CLAYTON JP- BY 5 X ATTOENE 7 Sept. 16, 1969 w, c Y JR v 3,467,067

RECIRCULATING TYPE ONCE-THROUGH STEAM GENERATOR Filed Dec. 27, 19s? 3 Sheets-Sheet 2 INVENTOR. WILLIAM H. CLAYTON J2- A T TOENE Y Sept E6, 1969 w. H. CLAYTON, JR 3,467,067

' RECIRCULATING TYPE ONCE-THROUGH STEAM GENERATOR Filed D60. 27. 1967 5 Sheets-Shem. 3

INVENTOR. WILLIAM H. CLAYTON JE- BY f/ ATTORNEY United States Patent US. Cl. 122-33 2 Claims ABSTRACT OF THE DISCLOSURE A once-through steam generator for operation at supercritical pressure and having recirculation through the furnace wall tubes. A heat exchanger located in the recirculating line is operative to transfer heat to incoming combustion supporting air. The heat exchanger may alternately be located in the mixed flow portion of the through-flow path at a location upstream of the furnace wall tubes. A recirculating pump is located downstream of the heat exchanger but upstream of the furnace wall tubes. Heat exchange fluid is circulated in a closed loop between a heat exchanger removing heat from the mixed flow, or recirculating line, and heat exchangers located in the combustion air supply. This heat exchanger for the air supply may be divided into two portions having one upstream of an air heater with the other downstream of the air heater. The relative effectiveness of each portion of this heat exchanger is regulated by bypassing either portion to effect air heater corrosion protection, maximum cooling of steam generator recirculated fluid, maximum air heating for combustion or maximum efficiency of a steam generator consistent with desired recirculated flow cooling.

Background of the invention This invention relates to once-through steam generators and in particular to steam generators with recirculation through the furnace wall tubes.

To start up a once-through steam generator a water flow of some sort must be established through the furnace wall tubes. While some units accomplish this by passing a through-flow in the order of 30 percent to waste, another method of accomplishing the flow is to recirculate water through the furnace wall tubes from the outlet to the inlet. This recirculation through the furnace Wall tubes may initially be either used as the total flow through the tubes or may be supplemented by a through-flow in the order of 5 percent of the designed full load through-flow. The nominal 5 percent through-flow is generally used to provide improved pressure control and generally uniform heating of various piping and headers outside of the recirculating flow path.

At the beginning, of startup, substantially cold Water is recirculated and a nominal through-flow of cold water is established. The unit is fired and as the temperature in the furnace wall tubes increases, the substantial recirculated flow is at high temperature while the nominal throughfiow remains at a generally low temperature. The mixed flow entering the furnace walls therefore is at high temperature and has a low density. At 3500 p.s.i. and 200 F. the density is 61.5 lb./ft. at 400 F., 56 lb./ft. at 600 F., 44 lb./ft. at 700 F., 33 lb./ft. at 750 F., 1b./ft. and at 800 F., 7 lb./ft. While the volumetric flow rate is constant, the mass flow rate is low due to this low density. The metal temperature of the furnace side of the furnace wall tubes is dependent not only on the rate of heat transfer from the furnace gases to the tube, but also the thermal resistance of the tube and the inside film between the tube and the water. Since this inside film Patented Sept. 16, 1969 is a function of the mass flow rate rather than the velocity or volumetric flow rate, poor heat transfer is obtained during this period of operation. While the furnace is operating at low loads and therefore relatively low heat absorption rates as compared to full load, local rates may exist in the general area of the burners. Furthermore, since the furnace absorbs most of its heat by radiant heat transfer, the heat transfer rate in the furnace does not decrease in proportion to load at these low loads and therefore is relatively high in proportion to the throughflow.

As the through-flow is increased and the recirculation is decreased, a greater quantity of the incoming cold water is mixed with a lesser quantity of the hot recirculated Water. The mixed water temperature therefore decreases and its density increases as load is increased on the steam generator. Therefore, safe operating conditions for the furnace wall tubes are easily achieved at the higher loads where there is a substantial through-flow, and safe conditions may also be easily achieved at the extremely low startup conditions where there is only a nominal heat input into the furnace. A critical condition still remains at intermediate loads between 5 and 30 percent of full load output where there is substantial firing in the furnace and a high quantity of recirculation which leads to high temperatures and low mass flows in the furnace wall tubes. Difficulties in this area have been generally overcome by overdesigning a recirculating system at other loads to cover this low load period of operation and by operating with high excess air in the furnace to minimize furnace absorption rates.

The extent of surface in the steam generator is dictated by full load heat transfer considerations, and therefore there is a substantial excess of surface during startup and low load operation. This excess surface leads to very low flue gas temperatures leaving the economizer and entering the air heater. Also in order to simplify control and provide furnace safety, a high quantity of excess air is used during startup which, in turn, also results in low gas temperatures entering the air heater. These low temperatures lead to problems with condensation and subsequent sulfur corrosion or plugging of the air heater. The problem has generally been solved by installing steamair heaters to preheat the air upstream of the air heater during startup or by bypassing air around the air heater. This low gas temperature also leads to a low air temperature leaving the air heater. This produces less effective combustion in the furnace and at times will be insuflicient for proper drying of coal in pulverizers during the early part of the startup. While the latter exists, it has not been found sufliciently severe to invest the money required to correct the problem.

Summary of the invention In my invention the recirculated fluid or the mixed fluid upstream of the furnace walls is cooled by the incoming combustion air supply, and the air supply is conversely heated. This cooling is accomplished upstream of the recirculation pump and therefore a substantial increase in the mass flow of water through the furnace wall tubes is achieved for low load operation when the density of the fluid entering the waterwall tubes of conventional recirculating type units tends to be low. In addition to cooling the recirculating fluid, the heating of the air may be accomplished upstream of the air heater for air heater corrosion protection and to minimize heat exchange surface requirements. It may be accomplished downstreamof the air heater to maximize the efficiency of the unit since this produces maximum utilization of the gas-toair air heater. Furthermore, the air heating may be selectively regulated so that it is accomplished as desired upstream and downstream of the air heater.

Brief description of the drawings FIGURE 1 is a schematic illustration of a steam generator using a closed loop heat exchanger between the steam generator recirculating line and the combustion air duct, and illustrating a recirculating pump in the mixed flow portion of the through-flow path;

FIGURE 2 is a schematic diagram similar to FIGURE 1 wherein the circulating pump is located in the recirculating line wherein the steam generator circulating pump is located in the recirculating line at a location downstream of the heat exchanger; and

FIGURE 3 is a schematic diagram of a steam generator wherein heat exchange is accomplished between the incoming combustion supporting air and the mixed flow portion of the through-flow circuit by passing this mixed flow portion through the air supply windbox.

Description of the preferred embodiments Feedwater pump 2 establishes a through-flow of water through the steam generator including economizer 3. The water passes then to the mixing vessel 4 and is conveyed to the furnace wall tubes 5 through a mixed flow conduit 7. These furnace wall tubes 5 are vertical tubes lining the walls of furnace 8 of the steam generator 9. This through-flow passes through the waterwall outlet line 10 at a temperature of 800 F., and the boiler throttle valve 12, thereafter passing serially through the low temperature superheater 13 and the high temperature superheater 14 where it is finally heated to a temperature of 1000 F. Steam is from here passed to a steam turbine driving an electric generator.

Circulating pump 15 floats on the through-flow system and operates to recirculate a portion of the through-flow through recirculating line 17 to the mixing vessel 4. The operation of this circulating pump to effect recirculation is more fully described in US. Patents 3,135,249 and 3,135,252. This recirculating line includes a check valve 18 which operates to prevent reverse flow through the recirculating line when circulating pump 15 is not op erating at a sufiicient head to effect recirculation.

Forced draft fan 22 establishes a flow of combustion supporting air through duct 23 to the furnace .windbox 24. This air passes through openings 25 in the furnace wall where it mixes with fuel passing through burners 27 to support combustion of the fuel within the furnace 8. The combustion products formed are passed upwardly through the furnace, passing horizontally through flue 28 over superheater 14, and vertically downward through flue 29 over super-heater 13 and economizer 3. The flue gases at a temperature of approximately 700 F. are then passed horizontally through flue 30 and the heated side of air heater 32. They then pass outwardly through flue 33 and discharge through a stack. The air heater 32 operates to cool the flue gases passing therethrough and simultaneously transfers this heat to the combusiton supporting air passing through duct 23.

During startup operation boiler throttle valve 12 is normally closed with the nominal through-flow passing through boiler extraction valve 37 to flash tank 38. From this flash tank water is drained through the water discharge line 39, while steam is supplied through steam admission line 40 to the superheater circuits. Excess steam may be spilled over to the condenser through spillover line 42. During normal operation at higher loads, valve 37 is closed with the entire through-flow passing through boiler throttle valve 12.

A heat exchanger 47 is located with the heating side 48 of this heat exchanger being a portion of the recirculating line 17. The heated side 49 of the heat exchanger forms a portion of a heat exchanger loop containing a heat exchange fluid such as pressurized water. Heat exchanger surface 50 is located in the air duct 23 at a location downstream of air heater 32 while heat exchange surface 52 is located in the air duct 23 at a location upstream of the air heater 32. Heat exchange fluid circulating pump 53 circulates fluid serially through the heat exchanger 47, the heat exchange surface 50, and the heat exchange surface 52. This heat exchange loop operates to transfer heat from the recirculating fluid which is at a temperature level of about 800 F. shortly after startup, to the incoming combustion supporting air which would normally be at a temperature of about F. upstream of the air heater and at a temperature of about 400 F. downstream of the air heater. Because of the low temperature air existing at heat exchange surface 52, less surface is required to transfer a given amount of heat. Therefore capital investment in heating surface is minimized to the extent that this location is chosen. However, the higher air temperature entering air heater 32 leads to a higher gas temperature leaving the air heater and therefore lower steam generator efiiciency. Since this is generally not a problem during loadoperation when my invention is most advantageous, a substantial amount of surface may be placed upstream of the air heater, if desired. If a higher efliciency is desired in the unit, more of the surface should be placed downstream of the air heater at heating surface 50. Heat exchange surface 50 may be made relatively ineffective by regulatingly opening bypass valve 54 while heat exchange surface 52 may be relatively less effective by opening bypass valve 55. Therefore, for any given surface installation of heating surfaces 50 and 52, the relative heat absorption may be regulated.

By heating the air with heating surface 52 upstream of the air heater 32, the average metal temperature of the air heater 32 surface is increased. This heating may be regulated to avoid the well known problemsof air heater corrosion due to condensation and formation of acid from the flue gas. The air heating accomplished by heat exchange surfaces 52 and 50 results in higher preheated air temperatures generally. This provides better drying if air is used for coal drying in a pulversized coal system at this time, and improved combustion. It further minimizes disturbance to circulation within the furnace wall tubes where this disturbance is due to localized cooling of certain air swept tubes. Such cooling can remove heat from the fluid within the tubes, increasing its density, with the weight of the water in these cooled tubes becoming such as to stagnate or reverse flow in the affected tubes. The higher air temperature minimizes'this detrimental cooling. 7

Since the fiow effected through the furnace walls is generally at a constant volumetric rate, a substantial increase in the mass flow through the furnace wall tubes can be effected by removing heat from the recirculated fluid when the recirculating pump is located downstream of the heat exchanger. In this 500,000 kilowatt'unit illustrated, an 11 percent through-flow rate amounts. to 400,000 pounds per hour. During startup operation the temperature is gradually increased until the furnace wall outlet temperature is about 800 F. At this time the economizer outlet temperature is 600 F. with a pressure of 3500 p.s.i. being maintained in the furnace wall system. The heat exchanger 46 removes million B.t.u.s per hour which corresponds to about a 3percent firing rate. Removal of this amount of heat by the heat exchanger, as illustrated, increases the mass velocity entering the furnace wall tubes approximately 32 percent. The installation of the heat exchanger 47 in the recirculating line 17 has several advantages. First, the high temperature level existing in this recirculating line minimizes the surface requirements for transferring the desired amount of heat. Furthermore, when the unit is operating at higher loads without recirculation, the heat exchange surface is removed from the through-flow path and therefore does not impose a pressure drop in the steam generating y The heat exchanger could however be located in the mixed flow line 7 upstream of the furnace wall tubes 8 and still be generally beneficial. At very low load operation where a high percentage of the flow is recirculated with a low through-flow, the temperature level in the mixed flow line 7 approaches that in the recirculating line 17. It is just this time when the mixed flow temperature is high and therefore of minimum density that the low mass flow problem occurs in the furnace wall tubes. The heat exchanger located in mixed flow line 7 therefore has suflicient temperature head to be effectively operative at these times of poor mass flow through the furnace wall tubes. As through-flow through the steam generator is increased and the recirculation is decreased, the temperature in mixed flow line 7 decreases. A heat exchanger located in this line would become ineffective at these higher through-flows, but since the temperature entering is already at a low level therefore presenting a dense fluid, there is no need for heat removal at this time. In either event maximum advantage of this heat exchanger can only be taken if the recirculating pump is located downstream of the heat exchanger 47.

In this respect FIG. 2 illustrates a steam generator arrangement substantially identical with that of FIG. 1 except for the location of the recirculating pump. The recirculating pump 65 is located directly in the recirculating line 17 but still at a location downstream of the heat exchanger 47. With the recirculating pump in this location a critical period of operation also occurs at high loads where it is difficult for the pump to generate sufficient head to effect recirculation due to the low density fluid passing through the pump. Therefore with this pump location, the location of heating surfaces 50 and 52 should be biased with more substantial heating surface in section 50. Since the heat exchanger is likely to be operated more consistently at higher load where efliciency of the unit is important, little use could be made of heating surface 52 without resulting in an increase in the temperature of the combustion products passing outward through duct 33 with a consequent decrease in steam generator efliciency.

FIG. 3 is a schematic illustration of a similar steam generator which does not use the intermediate heat exchange fluid. Mixing vessel 4 receiving cold water from economizer 3 and recirculated hot water through recirculating line 17 is located at an upper elevation. The mixed flow line '7 passes downwardly with at least a portion of this line passing through windbox 24. Since the portion within the windbox is heat exchange surface 67, this portion of the line may be comprised of a plurality of parallel tubes with extended heating surface, as desired. The circulating pump is again located in mixed flow line 7 at a location downstream of the heat exchange surface 67. This embodiment illustrates a simpler and less expensive arrangement of my invention. The heat exchange surface 67 inherently produces substantial heat transfer at low loads due to the low air temperature and high mixed flow temperature, while decreasing in effectiveness at higher loads. As previously discussed, it is during these low load periods that cooling of the recirculated fluid is required. While suffering the deficiency of substantial pressure drop at high loads if heat exchange surface 67 is used, a low pressure bypass may be inexpensively installed around heat exchange surface 67 for high load operation. Since the heat exchange surface 67 is exposed only to air flow, it will not be damaged due to the lack of flow therethrough.

While I have illustrated and described preferred embodiments of my invention it is to be understood that such is merely illustrative and not restrictive and that variations and modifications may be made therein without departing from the spirit and scope of the invention. I therefore do not wish to be limited to the precise details set forth but desire to avail myself of such changes as fall within the purview of my invention.

What I claim is:

1. A recirculating type through-flow steam generator comprising: a furnace; furnace wall tubes lining the walls of said furnace; a duct for conveying combustion supporting air to said furnace; means for establishing a flow of air through said duct; means for supplying fuel to said furnace and burning the fuel, creating a flow of combustion products; a mixing vessel; a superheater; a first conduit for conveying fluid from said mixing vessel to said furnace wall tubes; a second conduit for conveying fluid from said furnace wall tubes to said superheater; a third conduit for conveying fluid from said second conduit to said mixing vessel; means for establishing a through-flow serially through said mixing vessel, furnace wall tubes and superheater; means for recirculating a portion of the through-flow through said third conduit; and heat exchange means for transferring heat from the flow through one of said first and third conduits to the flow of air through said duct comprising a heat exchange fluid, first heat exchange means for transferring heat from the flow in one of said first and third conduits to the heat exchange fluid, second heat exchange means for transferring heat from the heat exchange fluid to the flow of air, and circulating means for recirculating the heat exchange fluid in a closed loop serially through said first and second means.

2. An apparatus as in claim 1 having also: air heater means for transferring heat from the flow of combustion products to the flow of air through said duct; said first heat exchange means transferring heat from the flow through said third conduit to the heat exchange fluid; said second heat exchange means transferring heat from said heat exchange fluid to said flow of air at a location upstream of said air heater; a third heat exchange means for transferring heat from said heat exchange fluid to said flow of air at a location downstream of said air heater means; said recirculating means recirculating said heat exchange fluid in a closed loop serially through said first, second and third heat exchangers; means for bypassing the heat exchange fluid around said second heat ex change means; and means for bypassing the heat exchange fluid around said third heat exchange means.

References Cited UNITED STATES PATENTS 2,170,345 8/1939 Bailey et al. 2,255,612 9/1941 Dickey. 2,883,832 4/1959 Arnow 122-1 XR 3,242,911 3/1966 Schroedter.

KENNETH W. SPRAGUE, Primary Examiner U.S. Cl. X.R. 122-406 

